In connection with refiners or grinders which have relatively rotatable refining members defining a refining gap, during passage of the raw material or pulp through that gap, considerable axial forces are produced which act against the refining members which have been set at desired refining gap distances therebetween. This invention thus precisely relates to a refiner of the rotating disc type for refining papermaking pulp and the like in which the raw material which is to be refined or treated in some way is passed through such a refining gap defined by a pair of axially adjustable refining discs rotating relative to each other. These refining discs thus rotate relative to each other in a plane which is perpendicular to the shafts associated therewith. Thus, at least one of the refining discs is axially adjustable and mounted on a rotary shaft, which, in turn, in response to the pressure acting on the disc itself, is axially movable with the adjustable refining disc. The pulp or raw material itself can be wooden chips, bagasse, fiber suspensions or similar-type materials which are fed to the central portion of the refining gap, through which this material is radially accelerated by the action of centrifugal force generated by rotation of the discs themselves. The thus-processed material is then discharged after the refining operation through a peripheral opening between the discs into a surrounding casing.
Axial movement of the rotary shaft must be controlled in order to maintain the predetermined refining gap between the refining discs. The size of the gap varies depending upon the intended use of the refiner. In conventional pulp refiners, for example, the gap normally has a dimension of between about 0.1 and 1 mm, while in refiners for waste paper the gap size can be as large as about 2.5 min. In connection with other applications, the refining gap may be as small as about 0.05 mm. Pulp refiners of this type are disclosed in U.S. Pat. Nos. 4,082,233; 4,253,233; 4,283,016; 4,378,092; and 4,801,099.
The rapid acceleration of material through the narrow refining gap creates axial pressure forces, which tend to separate the refining discs from each other, thereby widening the refining gap, as a result of which the efficiency of the refiner can be seriously deteriorated.
When the refiners or grinders are part of a closed or pressurized system, for example, where they are used in the treatment of a liquid slurry, additional forces must be supplied to the drive mechanism which are greater than the axial pressure forces acting upon the discs. This additional force is required not only to drive the discs in order to obtain the desired refining or grinding, but also to drive the discs against the liquid friction or hydraulic brake forces acting on the discs, so that additional axial load variations on the rotary shaft are obtained.
When the effect of these forces on the axial position of the rotary shaft are not effectively controlled, the refiner will break down. Furthermore, the resistance against these pressure forces increases substantially with increasing diameter of the discs.
Because of the increasing demand of refining systems having high capacities, which thus require large diameter refining discs, for example of the magnitude of 150 cm or greater, the absorption of these axial pressure forces has become a widely recognized problem.
Newly developed refiners have a disc diameter of between about 165 and 170 cm, a rotational speed of between about 1500 and 3600 rpm, and a power of between about 15,000 and 50,000 kW.
In order to better appreciate the enormous axial loads or pressure forces which are acting on the rotating shaft, one can imagine that a disc with a diameter of 150 cm rotating at 1,800 rpm develops a centrifugal force corresponding to about 2,800 g, which force then accelerates the material through the refining gap. This centrifugal force can apply an axial load of about 100 tons to the shaft itself, which load must therefore be taken up by the bearing structure. At a refining disc speed which is twice as high, i.e., of 3,600 rpm, the centrifugal force increases by a factor of four, according to Newton's law of force and motion. The centrifugal force thus increases to about 11,200 g, in which case the axial load on the rotating shaft can increase to about 200 to 400 tons. In connection with present bearing designs, such abnormally high axial loads must be distributed by use of a complicated bearing system, which requires a plurality of bearings and servomotors, with a resulting increase in the dimensions, as well as in the manufacturing costs of the refiner.
One example of a bearing structure of the aforesaid type is shown in U.S. Pat. No. 3,717,308, which discloses a bearing system with combined axial and radial bearings supporting the rotating shaft. Each of these bearings is coupled to a servomotor for taking up the axial pressure forces acting on the rotating shaft. Other examples of such bearing designs are shown in U.S. Pat. Nos. 4,118,800; 3,212,721; 4,073,422; and 3,276,701. U.S. Pat. No. 4,402,463 proposes yet another solution of the aforesaid problems.
The common feature of the state of art as set forth in the patents referred to above is that the hydraulic pistons in the servomotors for thrust bearings are non-rotary. U.S. Pat. No. 4,801,099 (Reinhall), however, proposes the use of one or more hydraulic rotational pistons, which are rigidly connected to the rotary shaft, and which entirely replace present systems with expensive and complicated axial, roller and/or block bearing systems. The bearing system utilizing rotating pistons according to Reinhall comprises one or more cylinder pistons mounted on the rotary shaft for rotating along with the shaft in a pressure chamber, which is formed in a stationary cylindrical housing, in which the piston(s) can be axially displaced. In this patent the axial piston displacement takes place by means of a pressure medium, which is supplied to at least one end of the piston(s) in a controlled manner, in order to continuously counteract the varying axial pressure forces acting on the movable rotary shaft, and in order to maintain the predetermined size of the refining gap.
This system, however, requires a plurality of sealing devices located at the inlets of the rotary shaft in the stationary cylindrical pressure housing as well as between the circumference of the rotating piston and the cylinder housing for its operation. These circumferential sealing devices are exposed to the vibrations of the rotating shaft which are caused by the bias of the refining elements and/or by the non-uniform distribution of the material over the surface of the refining elements. In order to prevent breakdowns, it is therefore necessary to maintain relatively large radial gaps at these sealing surfaces. The size of these gaps must therefore exceed the magnitude of the maximum radial vibrations. As a result, substantial amounts of the pressure medium supplied to the piston housing are lost as leakage through the radial sealing gaps. Furthermore, at the necessarily high relative hydraulic pressures of from about 100 to 400 bar, considerable energy and large, expensive pump installations are required.